EP2909502A1

EP2909502A1 - Torsional vibration damper assembly with pre-tensioning

Info

Publication number: EP2909502A1
Application number: EP13766074.2A
Authority: EP
Inventors: Thomas Dögel; Uwe Grossgebauer; Andreas Orlamünder
Original assignee: ZF Friedrichshafen AG
Current assignee: ZF Friedrichshafen AG
Priority date: 2012-10-17
Filing date: 2013-09-23
Publication date: 2015-08-26
Also published as: DE102012218926A1; US9964176B2; CN104736881A; CN104736881B; WO2014060191A1; US20160169321A1; DE102012218926B4

Abstract

Embodiments relate to a torsional vibration damper assembly (10; 10'), in particular for the drive train (100) of a vehicle, said assembly comprising a support assembly (12) that can be rotated about a rotational axis (A) and a deflection mass (34) that can move relative to the support assembly (12) in the peripheral direction. The support assembly (12) and the deflection mass (34) are rotatably coupled relative to one another by means of a plurality of elastic restoring elements (42) arranged in the peripheral direction and extending radially and each restoring element (42) can be deformed about a force application point (58; 58'; 60; 60') that is associated with said element (42) and can move in the radial direction under the action of centrifugal force. The assembly is characterised in that a first restoring element (42) is pre-tensioned in a first direction when the torsional vibration damper assembly (10) is immobile and that a second restoring element (42) is pre-tensioned in a second direction opposite to the first direction when the damper assembly is immobile.

Description

Rotary vibration damping arrangement with preload

Embodiments of the present invention generally relate to torsional vibration damping assemblies, preferably for the powertrain of a vehicle, and more particularly to torsional vibration damping assemblies having alternately bi-directionally biased elastic return members.

In order to dampen vibrations, in particular torsional vibrations caused, for example, by rotating components (such as a crankshaft) in a motor vehicle, numerous concepts are known. In addition to balance shafts so-called torsional vibration dampers can be used additionally or alternatively. Such torsional vibration dampers generally include damping or deflection masses, by the inertia of which undesirable torsional vibrations can be damped. A well-known torque-transmitting torsional vibration damping concept, for example, to decouple the flywheel mass system of the engine from the transmission and drive train is z. B. the dual-mass flywheel with a primary flywheel, a secondary flywheel and a mounted therebetween torsional vibration damping arrangement.

From DE 10 2010 053 542 A1 a torsional vibration damping arrangement or a Tilger is known in which Auslenkungsmassenpendeleinheiten arranged around a support around, by means of a plurality of fixed thereto and radially inwardly extending elastically deformable return elements (eg leaf springs) with respect the carrier supported in the circumferential direction comprises deflection mass. In the carrier centrifugal weights or support elements are provided radially displaceable, on which the radially inwardly extending restoring elements in the circumferential direction can be supported on respective Trägerabstützbereiche or force application points. The support elements are biased by this associated and on the Auslenkungsmasse supporting biasing springs radially inward into a base position. If the centrifugal force load is low or nonexistent, the centrifugal weights or supporting elements are held under the prestressing action in the base layer. With increasing speed, the support elements displace due to centrifugal force under increasing compression of the biasing springs radially outward, causing the Trägerabstützbereiche, at which the radially inwardly of the Auslenkungsmasse extending restoring elements can be supported, are displaced radially outward Shen. This changes the available for deflection free length of the return elements between their connection to the deflection mass and the respective Trägerabstützbereichen in which they are supported on the support elements with respect to the carrier in the circumferential direction. This variation of the free length thus also influences the effective pendulum length, the shortening of which leads to an increase in the natural frequency of the deflection mass pendulum units. As a result, the stiffness and thus also the natural frequency of the deflection mass pendulum units can be varied as a function of rotational speed, in a sense that the stiffness and thus also the natural frequency of the torsional vibration damping arrangement increases with increasing rotational speed. This is intended to attempt to achieve a speed adaptation of the deflection mass pendulum units to a vibration excitation order.

Known torsional vibration damping arrangements thus have an adjustment system which, as a function of the rotational speed, detunes the natural frequency of the torsional vibration damping arrangement or of the absorber, in order thus to purposely eliminate a vibration excitation order over a wide rotational speed range. The adjustment system consists preferably of a plurality of centrifugal weights or support elements, which are distributed symmetrically on the circumference of the carrier to minimize imbalance, and acts on the centrifugal force under speed. Furthermore, the adjusting system comprises at least one restoring element or an adjusting spring, which exerts a restoring force radially inward on the flyweight. The centrifugal force of the flyweights and the restoring forces of the springs are coordinated so that a desired position of the flyweight depending on the applied speed is adjusted (order tracking). The position of a centrifugal weight determines the force application point or pendulum point on a return element (eg bending spring or Tilgerfeder) and thus directly affects the stiffness and thus the natural frequency of the absorber. By circumferential clearance (ie play in the circumferential direction) between the restoring element and the force application point or pendulum point, the stiffness characteristic of the absorber can be influenced. It is an object of the present invention to provide a linear or strictly monotonous course of the stiffness characteristic curve over an entire vibration width or an entire vibration angle range for a respective rotational speed.

This object is achieved by a torsional vibration damping arrangement and a drive train for a vehicle according to the independent patent claims.

Friction of the flyweights on the return elements that essentially exert tangential forces influences the radial centrifugal force-dependent adjustment of the flyweights along the return elements (for example leaf springs). In order to minimize the influence of friction, it is expedient to provide a part of the oscillation angle of the torsional vibration damping arrangement in the load-free state for the radial adjustment of the flyweights. Although circumferential play between flyweights and restoring elements (Tilgerfedern) may be useful for this purpose, however, is in the trade-off with a constant as possible Tilgersteifigkeit with the most linear and strictly monotone characteristic curve for the respective speed. Embodiments deal with the solution of the problem described.

According to a first aspect of the present invention, a torsional vibration damping arrangement is provided, which can be used in particular for damping torsional vibrations in the drive train of a motor vehicle, such as a combustion-powered and / or electrically driven vehicle. The torsional vibration damping arrangement comprises a carrier arrangement rotatable about a rotation axis and a pendulum mass which is relatively movable relative to the carrier arrangement in the circumferential direction (ie tangentially). The carrier arrangement and the deflection mass are rotatably coupled relative to each other via a plurality of circumferentially arranged and substantially or substantially radially extending elastic return elements, which may also be referred to as Tilgerfedern or pendulum rods. In this case, at least one restoring element, but in particular all restoring elements, in each case movable or pendulum-type about a force application point which is movable under centrifugal force in the radial direction and which is assigned to the restoring element. In other words, the restoring element is one at each Fliehkrafteinwirkung in the radial direction movable and the restoring element associated force application point deformable or bendable. To achieve the above object, it is proposed to bias a first elastic restoring element in the rest position of the torsional vibration damping arrangement in a first direction and to bias or fix a second elastic restoring element in the rest position in a second direction opposite to the first direction. In this case, the restoring elements can be prestressed or arcuate in different or opposite directions with respect to a respective associated point of force application. In the present case rest position means a state of undeflected deflection mass.

In some embodiments, the first and second return elements may form a pair of circumferentially immediately adjacent or oppositely disposed and circumferentially elastically deformable return elements. In this case, the torsional vibration damping arrangement may comprise a plurality of circumferentially arranged pairs of return elements or corresponding Auslenkungsmassenpendeleinheiten.

According to exemplary embodiments, an elastic restoring element of the torsional vibration damping arrangement can comprise, for example, in each case a restoring spring, in particular a leaf spring or a bar spring, in particular with a linear force characteristic. In this case, the restoring element can be fixed or fixed with respect to the deflection mass and / or with respect to the carrier arrangement. In some embodiments, a radially extending return spring is fixed to the Auslenkungsmasse and engages from there into a corresponding (radially further inward) guide the carrier assembly, in which also the return spring associated centrifugal weight can move radially up and down and thus a radially variable Kraftangriffsoder Pendulum point for the return element provides. The restoring element is elastically deformed or bent in the circumferential direction upon deflection of the deflection mass in accordance with a force acting on the force application point in the circumferential direction.

In embodiments, the bias (in the circumferential direction) of the return elements should be large enough, for example, greater than 0.5% or 1% of the maximum deflection of a return element to the circumferential clearance in Verstellsys- Tem of the force application points to eliminate any manufacturing tolerances. The bias of the return elements, ie the preload, sets a maximum allowable deflection (maximum allowable swing angle) of the return elements due to maximum allowable bending stresses. Thus, the swing angle of the return elements is further limited to avoid overloading and thus damage to the restoring elements or the Tilgerfedern.

The bias voltage or the Vorspannweg should not exceed a certain upper limit according to embodiments. This upper limit depends on the radial displacement of the flyweights and the expected relative swing angle of the deflection mass (in the circumferential direction). The same circumferential clearance leads radially au Shen to a smaller clearance angle than radially inside, as well as the proportion of the swing angle under bias radially outward smaller than radially inside. At high speeds, the stiffness of the restoring elements or Tilgerfedern is increased, whereby their deflection decreases assuming a constant Tilgermoments against lower speeds. If the preload of the Tilgerfedern reaches a value that is greater than the swing angle (for example, at high speeds), no load-free state is reached, whereby the adjustment of the flyweights subject to additional frictional forces.

According to some embodiments, the bias of the restoring elements in the rest position may be in a range of less than 10%, preferably less than 5%, of the maximum deflection of the restoring elements. In the rest position, a biasing force acts in the circumferential direction on the restoring elements, which causes a (small) deflection of the restoring elements.

In embodiments, the opposing biasing forces resulting from the opposing biases of the elastic return elements in the rest position of the torsional vibration damping arrangement, ie in the state undeflected by a base position, are equal in magnitude. In other words, the opposite biasing forces complement each other in the rest position of the torsional vibration damping arrangement to zero, so that there is no deflection without external torque action. In embodiments of a torsional vibration damping arrangement, at least one (radially) movable force application point is provided for each restoring element by a centrifugal weight which can be moved or displaced along the associated restoring element in the radial direction under centrifugal force, which is also referred to herein as a support element. For this purpose, the centrifugal weight can have circumferential support areas, for example in the form of bolts, either on one side or on both sides of the restoring element, about which the restoring element can oscillate or deform when deflected. According to embodiments, the first and the second elastic return element in the rest position with at least one of their respectively associated force application points, ie a circumferential support area, are in direct pressure contact to obtain the respective rest position bias of the return element. Preferably, a restoring element in the rest position with only exactly one force application point (eg bolt of the centrifugal weight) in direct contact.

According to embodiments, the first and the second restoring element form a pair of restoring elements, which mutually neutralize each other in the rest position of the torsional vibration damping arrangement. A along the first return element in the radial direction movable first flyweight and along the second return element in the radial direction movable second flyweight standing with their respective associated first and second return elements on each different sides of the return element in contact to the pair return elements in the rest position in to bias opposite directions. If the restoring elements are designed in the form of leaf springs, deflections of the leaf springs in different directions result in the rest position.

In embodiments of the present invention, the position of the at least one movable force application point on the flyweight with respect to the associated restoring element in the rest position of the torsional vibration damping arrangement can be asymmetrical, ie not mirror or axisymmetric. This can be achieved, for example, by virtue of the fact that each return element of the torsional vibration damping arrangement is only exactly one movable in the radial direction Force application point is assigned in or at a flyweight. In this case, the respective force application points of the first and the second restoring element of the pair can be arranged on respectively different sides of the associated restoring elements. Thus, while for a first return element of the force application point is in the counterclockwise direction, the force application point of the second return element is arranged in a clockwise direction with respect to the second return element. The first and second return elements associated Auslenkungsmassen- pendulum units can be arranged according to embodiments in the circumferential direction alternately or opposite, so that the different biasing forces complement each other in the rest position to zero.

In other embodiments, each return element can be assigned two radially movable force application points, wherein the two force application points under centrifugal force can move radially up and down on different sides of the respective return element and wherein the two force application points are arranged asymmetrically with respect to the respective restoring element in the rest position , This means, for example, that the restoring element is on one side in the rest position in direct contact with a first point of force application, while on the other side between the restoring element and a second point of force application in the rest position a game in the circumferential direction (circumferential clearance) exists. The two force application points are thus at different distances to the restoring element, wherein one distance is zero, the other is greater than zero. In such embodiments, it may further be provided that the asymmetrical arrangement of two first force application points with respect to the first return element is inverse to the asymmetrical arrangement of two second force application points with respect to the second return element. A structure of the adjustment system of the torsional vibration damping arrangement with such biassed mutually actuated restoring elements can, compared to other possible embodiments, with the same number of return elements best approximate an ideal (backlash) stiffness characteristic.

By asymmetric arrangement of the force application points formed by means of bolt elements in flyweight in the way that in zero position (rest position) a deflection or deflection of the return element takes place, or optionally by asymmetric arrangement of the return elements relative to the centrifugal weight or by asymmetric arrangement of flyweights (guideway of flyweights) to the reset elements, or a combination of the various possibilities, a mutual tension of the return elements can be realized and tolerance-related game in the system are eliminated.

By an asymmetric design of the flyweight, the remainder of the damper or torsional vibration damper assembly, such as the damper, can be used. a clamping of the return elements, as well as guideways of centrifugal weights are symmetrical and only by mutual installation of the reverse asymmetric flyweights mutual bias can be generated. Here, by constructive measures, such as. Compensation holes, the center of gravity of the centrifugal weight can be influenced as low as possible for the radial adjustment under speed.

In embodiments of the torsional vibration damping arrangement, therefore, the first restoring element and the second restoring element are arranged at least one, preferably exactly one, of their respective associated points of application of force without circumferential movement play. In other words, there is a direct contact between restoring element and at least one point of force application, which is formed for example by a Umfangsabstützbereich a flyweight.

In accordance with another aspect of the present invention, a powertrain for a vehicle is provided that includes at least one torsional vibration damping assembly according to embodiments.

Further advantageous embodiments and further developments are the subject of the dependent claims and the following detailed description.

Embodiments of the present invention will be explained below with reference to the accompanying figures. Show it: 1 shows a longitudinal sectional view of a torsional vibration damping arrangement;

Fig. 2 is an axial view of the torsional vibration damping arrangement of

Fig. 1 in the viewing direction I I in Fig. 1;

Fig. 3 is a representation corresponding to Figure 2, wherein a support plate of a ring-like Auslenkungsmasse is omitted ..;

4 shows in its illustrations a) and b) a support of the torsional vibration damping arrangement of FIG. 1 in a perspective view, viewed from different sides;

Fig. 5 in your representations a) and b) a ring-shaped deflection mass in longitudinal section, cut into different

Cutting planes;

6 is a perspective view of the annular deflection mass;

Fig. 7 is a detail view of a deflection mass pendulum unit;

8 is a view of a support element of the Auslenkungsmassenpendeleinheit, viewed from radially au Shen;

FIG. 9 the supporting element of FIG. 8 in a perspective view; FIG.

Fig. 1 0, the support member of Figure 8 in side view.

Fig. 1 1, the support member of FIG. 8, taken along a line Xl-Xl in

Fig. 1 0; the periodic deflection of a deflection mass of the torsional vibration damping arrangement with bilateral support of restoring elements; a deflection mass pendulum unit with double-sided, play-related arrangement of force application points around a return element; a deflection mass pendulum unit with mutual, play-free arrangement of force application points around a restoring element a comparison of stiffness characteristics for backlash-free and backlash-dependent mutual drive according to FIGS. 13 and 14; a representation corresponding to FIG. 3 of a modified in particular in the region of the support elements Ausgestaltungsart; an enlarged detail view of a deflection mass pendulum unit of the torsional vibration damping arrangement of Fig. 16; a support member of the deflection mass pendulum unit of Figure 17, viewed from radially outside; the support member of Figure 18 in a perspective view. the support member of Figure 18 in side view. the support member of Fig. 18, taken along a line XVIII-XVII in Fig. 20; the periodic deflection of a deflection mass with one-sided support of restoring elements; a pair of unilaterally actuated and biased return elements; a pair of mutually actuated and biased return elements; a comparison of stiffness characteristics for mutual and one-sided control of reset elements; a comparison of stiffness characteristics for mutual AnSteuerung of return elements, backlash-free and playful but mutually biased; a comparison of stiffness characteristics for mutual and one-sided control of reset elements; in a schematic representation of a drive train for a vehicle with a torsional vibration damping arrangement constructed according to the invention; a representation corresponding to FIG. 28 of an alternative embodiment; a representation corresponding to FIG. 28 of an alternative embodiment; a representation corresponding to FIG. 28 of an alternative embodiment; a partial longitudinal sectional view of a hydrodynamic torque converter with an integrated therein torsional vibration damping arrangement with inventive design. In the following description, like reference numerals designate functionally identical or similar components or parts.

1 to 3 show a generally designated 10 torsional vibration damping arrangement, which can be integrated to fulfill the functionality of a speed-adaptive absorber in a drive train of a vehicle or coupled to this. The torsional vibration damping arrangement 10 comprises a carrier arrangement 12 to be fixed by screwing to a drive train component for common rotation therewith about an axis of rotation A. In this carrier arrangement or this carrier 12, in the representations of FIGS. 3 and 4, at approximately approximately circumferential positions uniform circumferential distance guides 14 are provided, in which flyweights, which are also referred to below as support elements 1 6, are received radially movable. The guides 14 are formed as substantially radially extending, slot-like recesses which are bounded radially inwardly by stops 18 which define a radially inner base position of the flyweights 1 6. The flyweights 1 6 can be held biased by bias springs formed as helical compression springs 20 radially inward to rest against the stops 18, so in their base position. In this case, the biasing springs 20 can be supported on a radially outer annular edge region 22 of the carrier 12.

On the support 12, a support disk 28 is basically rotatably supported about the axis of rotation A with respect to the support 12 via a radial bearing 24 and a thrust bearing 26. In its radially outer region, the carrier disk 28 can carry a ground ring 30, for example by screwing on one axial side. On the other axial side of the carrier disk 28, for example, a further ground ring 32 may be defined. The support disk 28, together with the ground ring 30 and possibly also the ground ring 32, a deflection or pendulum mass generally designated 34, which in the circumferential direction (ie, tangential direction) can oscillate about the support assembly 12 to damp torsional vibrations. By a plurality of also visible in FIG. 6, circumferentially elongated recesses 36 and an axial securing ring 38 on the side facing away from the carrier 12 side of the support plate 28 holding bolts 40, such as bolts, the Trägerschei- be 28 and thus the deflection mass 34 axially secured to the carrier 12. Due to the circumferential movement play of the bolts 40 in the recesses 36 of the support plate 28, the deflection mass 34 is rotatable in a corresponding circumferential movement clearance with respect to the support 12 about the axis of rotation A, so that by relative cooperation of the bolt 40 with the recesses 36, a relative rotation angle limit is provided.

The deflection mass arrangement 34 is coupled to the carrier 12 by a plurality of circumferentially successive, substantially radially extending springs or restoring elements 42 for transmitting power. These restoring elements 42, which are designed here as leaf springs or generally as bending beams, can be fixed in the radially outer area 44 by a respective clamping arrangement 46 on the grounding ring 30. Starting from this definition, they extend radially inwardly through openings 48 in the edge region 22 of the carrier assembly 12 into a respective biasing spring 20 for radially biasing the flyweights 1 6 inside.

As FIG. 7 illustrates, the or each restoring element 42 extends with its radially inner end region 50 into or through a central opening 52 of an associated centrifugal weight or flyweight 16. In the region of the opening 52, two circumferential support areas 58, 60, provided for example on pins or bolts 54, 56 are provided on the flyweight 16, which represent radially displaceable force application points about which an associated return element 42 can deform. Here, the mass 34 oscillates about the axis of rotation A. These Umfangsabstützbereiche or force application points 58, 60, which may lie in the circumferential direction in some embodiments, both sides of the radially inner end portion 50 of the associated return element 42, but asymmetrically, define in their entirety a Trägerabstützbereich 62nd whereas in the region in which the radially outer end region 44 of the restoring element 42 is fixed to the mass ring 32 or, more generally, the deflection mass 34, a deflection mass support region 64 is formed. As will be explained below, the return element 42 between the two points of application of force or abutments 58, 60 can be accommodated at least on one side without movement with preload to allow occurring under centrifugal force radial radial movement of the support member 16 in the associated guide 14 in the carrier 12. In order to protect the biasing spring 20, the support member 16 may have two axially oriented side guide projections 66, 68 which extend into associated, substantially radially extending guide recesses 70 of the carrier 12 and 71 of the carrier disc 28 and guided therein radially movable or can be included and form a radial stop. In order not to impair their relative rotation with respect to the carrier 12, in particular by the interaction of the guide projection 68 with the carrier disk 28, the recesses 71 can have a greater circumferential width than the recesses 70 in the carrier 12. Further, tilting of a supporting element 1 6 occurring under centrifugal force can thereby occur be prevented that its center of gravity M is approximately centrally located in the opening 52.

In the above with reference to FIGS. 1 to 1 1 explained with respect to their structural design torsional vibration damping arrangement 10 respectively forms a radially movable in the carrier 12 flyweight or support member 16, the cooperating with this restoring element or the Tilgerfeder 42, the flyweight. 1 6 radially inward in its recognizable in Fig. 7 radial base position (at rest of the torsional vibration damping arrangement without deflection in the circumferential direction) biasing biasing spring 20 and the Auslenkungsmasse 34 each kenkungsmassenpendeleinheit 72. In the illustrated embodiment, only a total of ten such Auslenkungsmassenpendeleinheiten 72th provided, wherein the carrier 12 is a common carrier for the support elements 16 of all Auslenkungsmassenpendeleinheiten 72 and the Auslenkungsmasse 34 is a common Auslenkungsmasse for all Auslenkungsmassenpendeleinheiten 72. However, the principles of the present invention could in principle also be realized if a separate carrier is provided in association with each or at least part of the deflection mass pendulum units 72 and / or if a separate one is assigned to all or part of the deflection mass pendulum units 72 Displacement mass is provided. For the sake of stability and in order to avoid undesired vibration states or to obtain a synchronous vibration behavior of all deflection mass pendulum units 72, however, at least the combination of all the deflection masses into a common, annular deflection mass 34 is advantageous.

FIG. 12 shows the periodic oscillation behavior of the pendulum or deflection mass 34 plotted against time or an introduced (rotational) oscillation, generally designated in FIG. 12 as an absorber deflection. In the area of the zero crossing, a grayed out deflection range can be recognized. This is the area which would be obtained by a playful receiving the elastic return elements 42 between the associated Umfangsabstützbereichen 58, 60 and bolts 54, 56, as shown playfully in Fig. 13.

With periodic deflection of the deflection mass 34 and corresponding periodic back and forth deformation, ie deflection, the return elements 42 in the circumferential direction would always in the zero crossing of the absorber deflection, ie in the basic relative position between the carrier 12 and Auslenkungsmasse 34, in which the return elements 42 due to the circumferential clearance are not curious, caused by the playful recording of the radially inner end portion 50 of the return elements 42 between the Umfangsabstützbereichen 58, 60 briefly a state without power transmission between the carrier 12 and the deflection mass 34. In this state, the restoring elements 42 do not stress the associated supporting elements 16 due to the play or the air between the restoring element 42 and the points of application of force 58, 60 in the circumferential direction, so that a radial displacement of the supporting elements, or substantially unimpaired, of such loads and thereby caused friction effects Slide blocks 1 6 may occur. This displacement of the sliding blocks (or centrifugal weights) 1 6, which is as unimpaired as possible by the effects of friction, can be fundamentally assisted by the fact that the torsional vibration damping arrangement 10 is supported in a fluid such as a fluid. B. oil, filled or fillable housing is added. This has the consequence that the support elements 1 6 experience a lubricating effect with respect to the carrier 12 and also the associated return elements 42 and thus can be moved easily under centrifugal load. The deflection mass pendulum unit 72 shown in FIG. 13 thus has a constructive peripheral clearance between the restoring element 42 and the bolt cooperating therewith or otherwise designed linear pick-off elements 54, 56 of the support element 1 6. The corresponding restoring element characteristic or spring characteristic curve is shown in FIG. Curve b) shown. In this case, the (circumferential) game expresses itself in the form of a force-free state in the area 82 of the system change of the restoring elements 42. In this area 82, the supporting elements or centrifugal weights 1 6 can negatively adjust neglecting frictional forces, virtually without friction. Advantageous is this game variant with regard to simplicity of construction (equal parts, no forced orientation during assembly). On the one hand, the clearance in the adjustment system must be selected to be large enough in accordance with the manufacturing tolerance chain in order to prevent any jamming of the displaceable parts in any case. On the other hand, high process capabilities of the manufacturing machines and a check of the adjustability are necessary to keep the game as low as possible.

The circumferential play should be as minimal as possible in order to influence the Tilgersteifigkeit as little as possible, since a sum stiffness is reduced by play, which has a big impact, especially at small angles of oscillation. In other words, from this perspective of the sum rigidity, it is desirable to adopt a construction of the deflection mass pendulum unit 72 shown in FIG. 14, which is the ideal state (no play) in driving the bending springs 42. The associated characteristic or spring characteristic is shown in FIG. 15, curve a). However, this backlash-free state can not or only with difficulty (process capability) or only be implemented by additional measures (mechanisms for tolerance compensation) from a manufacturing point of view. The ideal backlash-free control according to FIG. 14 also conflicts with the ease of the adjustment system of the deflection mass pendulum unit 72. In this context, play-free always means a certain amount of friction which negatively influences the reaction speed of the adjustment system when the speed changes. For an adjustment of the centrifugal weights 1 6 under centrifugal force with negligible frictional forces theoretically only the load-free states are available, which are available in this case only at exactly the punctual moment 84 of the investment change of the bending spring 42 from train to push or vice versa. An asymmetrically modified embodiment, in particular with respect to the structure of the supporting elements or sliding blocks 16, is illustrated in FIGS. 16 to 22 and will be explained with reference to these figures. The basic structure corresponds to the above-described, so that reference can be made to the relevant statements. It can be seen in particular in FIGS. 17 to 21 that, with the support elements or sliding blocks or flyweights 1 6 'shown here, only one pin / pin 54' or 56 'can be provided and accordingly only one force application point or pendulum point 58', 60 'may be provided on Trägerabstützbereich 62'(-> asymmetry). The return element 42 can thus be supported only in a circumferential direction in the Trägerabstützbereich 62 '. As shown in FIG. 16, the overall structure may be such that circumferentially successive deflection mass units 72 'alternately have a pin 54' for providing a point of application of force in a first circumferential direction and the following deflection mass pendulum unit 72 'A pin 56' is provided in order to realize there a force application point in a second, opposite circumferential direction, ie on the other side of the return element. As a result, only half of the return elements of all the deflection mass pendulum units 72 'are effective at each half oscillation, which halves the overall rigidity of the torsional vibration damping arrangement 10'.

It should be noted that, of course, the circumferential sequence of such different deflection mass pendulum units 72 'does not necessarily have to be alternating, as shown in FIG. 16. Also, in each case a plurality of deflection mass pendulum units 72 'with basically the same construction, ie support functionality in the same circumferential direction, could be arranged consecutively. For reasons of symmetry and to avoid imbalance, however, the alternating arrangement shown in FIG. 16 is particularly advantageous.

In order to avoid centrifugal force-induced tilting of the supporting elements 1 6 'in this embodiment as well, the center of mass M can advantageously also be located centrally in the opening 52. In order to achieve this, in order to compensate for the additional mass of a respective pin 54 'or 56' in the support Lement 1 6 'one or more holes or openings 74 and optionally inserted therein mass elements may be provided (see Fig. 21).

In Fig. 22, the operation of a thus constructed Auslenkungsmassen- pendulum unit 72 'with mutual actuation of the return elements 42 is illustrated. It can be seen that a deflection mass pendulum unit 72 'only acts during half oscillations for generating a restoring force. By cooperation of a plurality of then differently designed deflection mass pendulum units 72 ', i. with force application points on different sides, in each case a restoring force functionality can be achieved over the entire oscillation course. Further, one recognizes also in Fig. 22 a gray background area near the zero crossing of the absorber rash. In principle, it could also be provided here that the respectively present pin or point of force application has a slight circumferential distance (i.e., circumferential clearance) to the associated restoring element 42 in the zero crossing or idle state. Since, however, whenever a restoring element 42 is in a non-effective oscillation phase with respect to a deflection mass pendulum unit 72 ', the restoring element 42 with its radially inner end region 50 stands out from the only existing circumferential support region, according to embodiments such a backlash can be dispensed with it is shown in FIG. 23.

FIG. 23 shows a pair of two complementary deflection mass pinning units 72 'according to an exemplary embodiment. In the pair, a first return element 42 (eg left) is biased in a first direction in the rest position of the torsional vibration damping arrangement 10 (or at zero crossing of the absorber deflection) and a second restoring element 42 (eg on the right) in the rest position in a second direction opposite to the first direction Direction biased as indicated by the differently oriented arrows. In this case, the different biases or deflections of the two restoring elements 42 of the pair are preferably selected such that opposing biasing forces resulting from the opposite preloads in the rest position or the zero crossing are equal in magnitude. It can be seen that the first return element or the first bending spring 42 (left) in the rest position of the torsional vibration damping assembly 10 'in the first direction by direct contact to a first return element 42 associated force application point 58' (on the bolt 54 ') is biased and in that the second restoring element 42 (right) in the rest position (ie no deflection) is biased in the second direction by direct contact with a force application point 60 '(on pin 56') associated with the second restoring element 42. In other words, the two restoring elements 42, which form a pair of associated restoring elements, are also directly in the rest position at their respective force application points 58 ^' and 60'. As a result, according to embodiments in the rest position of the torsional vibration damping assembly 10, a respective bias or deflection of the return elements 42 in a range of 1% to 10% of a maximum deflection of a return element 42, such as a leaf spring result.

The embodiment shown in FIG. 23 shows a first centrifugal weight 1 6 'that can be moved along the first restoring element 42 (left) in the radial direction and a second centrifugal weight 1 6' that can be moved along the second restoring element 42 (right) in the radial direction. The flyweights or sliding blocks 1 6 'are connected to the first and the second restoring element 42 on respectively different sides, namely once left or counterclockwise (contact point 58') and right or clockwise (contact point 60 ^' ), the return element 42nd in direct pressure contact to bias the pair of return elements 42 in the rest position of the torsional vibration damping assembly 10 'in opposite directions. In the embodiment of FIG. 23, each of the two return elements 42 is associated with exactly one in the radial direction movable force application point, 58 ^' (left) and 60' (right). In this case, the respective force application points 58 ', 60' of the first and the second restoring element 42 are arranged on respectively different sides of the restoring elements 42 in order to obtain the opposite preloads.

FIG. 23 shows a one-sided actuation of the restoring elements or leaf springs 42. Preferably, in this embodiment, a pairwise mutual arrangement of the restoring elements 42 is used. That is, the first and second returns Control element 42 form a pair of restoring elements or deflection mass pendulum units 72 'arranged in the circumferential direction next to one another or opposite one another. In this case, the torsional vibration damping arrangement 10 'has a plurality of circumferentially arranged pairs. By way of example, FIG. 16 shows 5 pairs of deflection mass pendulum units 72 ', each with mutual or opposite bias, so that the total bias voltages are zero.

If the restoring elements 42 are prestressed against one another or oppositely, assuming the same number of restoring elements 42, a Tilger characteristic results, as shown in FIG. 25, curve c). In the pretensioning region 86, an ideal characteristic curve (see curve a), which drops to half the stiffness at restoring element bending greater than the preload, since again only half of all restoring elements 42 of the torsional vibration damping arrangement 10 'act in this case. In the area outside the bias, the adjustment of the return elements 42, which are no longer engaged, adjust without load. Fig. 25, curve c-1 .5 and c-2.0, represent the course of the absorber characteristic curve for each of the 1 .5-fold and 2.0-fold number of return elements 42 compared with curve c).

The differently biased return elements 42 are subjected to a swelling load in the arrangement according to FIG. 1 6 or FIG. 23 and are therefore characterized with regard to the service life compared to an alternating load, as described for example with reference to FIGS. 13 and 14. better off. As a result, in comparison to the alternating load with two force application points 58, 60 per return element 42, higher stresses or deflections and thus larger oscillation angles can be permitted. If the tuning (spring stiffness) of the absorber is designed so that on the other hand half a number of effective return elements 42 is sufficient, or the return elements 42 are correspondingly different dimensions, so that half the effective number of springs provides the same rigidity, as in an arrangement shown in FIG. 14 (compare FIG. 23, curve a) as ideal rigidity), the sum stiffness is twice as great for small angles of oscillation (smaller than or equal to the preloading angle) and for angles of oscillation greater than the preloading angle decreases with increasing angles in the direction of design rigidity from. The design stiffness is theoretically never achieved, but always slightly exceeded. Of advantage can this effect at higher speeds, if the swing angle are small and a radial end of the adjustment range of the flyweights 1 6 is reached, the order can be kept for a certain speed range before it falls below the tuning order. At low speeds, the excitation moments are the largest according to experience, whereby the torsional vibration damping assembly 10 'performs large swing angle and the sum stiffness thus corresponds approximately to the design. Since there is no ideal restraint for the restoring elements 42, the theoretical rigidity of the restoring elements 42 will practically never be achieved, but always somewhat lower. Since in the present embodiment, the sum stiffness is always slightly above the theoretically designed stiffness, this can have a compensating effect.

While the embodiments described with reference to FIGS. 16 to 23 are characterized in that each restoring element 42 of the torsional vibration assembly 10 'is assigned exactly one radially movable force application point 58' or 60 ', wherein the respective force application points of the first and the second restoring element 42 on are each arranged on different sides of the return elements 42 in order to obtain the opposite biases, further embodiments also provide other asymmetrical arrangements of force application points about a return element 42.

FIG. 24 shows by way of example a further embodiment in which each restoring element 42 of a torsional vibration arrangement 10 is assigned two force application points 58, 60 which are movable in the radial direction. Here again, a first restoring element 42 (eg on the left) is biased in a first direction in the rest position of the torsional vibration damping arrangement 10, while a second restoring element 42 (on the right) is biased in a rest position in a second direction opposite to the first direction (see arrows). , Again, the opposite bias can again be achieved by direct contact of the return element 42 to one of its two associated force application points 58, 60.

As with the examples described with reference to FIGS. 1 to 14, here the two points of application 58, 60 under centrifugal force move in opposite directions. Gend on different sides (left, right) of the respective return element 42 radially outward or inward. However, unlike the torsional vibration assemblies 10 described above, the two two force application points 58, 60 of FIG. 24 are asymmetrically disposed with respect to their respective return member 42 in the rest position (ie, no deflection of the deflection mass 34) to obtain the opposite biases. In other words, in the left deflection mass pendulum unit 72 of the pair of deflection mass pendulum units 72 shown in FIG. 24, the restoring member 42 in the rest position is in direct contact with the left pin 54 and the load point 58 of the left flyweight 1 6 between the left return member 42 and the right bolt 56 or force application point 60 of the left flyweight 1 6, however, there is circumferential play. In the right Auslenkungsmassenpendeleinheit 72 behaves exactly the reverse, that is, the right restoring element 42 is in rest position in direct contact with the right pin 56 and force application point 60 of the right flyweight 1 6. Between the right restoring element 42 and the left pin 54 and Force application point 58 of the right flyweight 1 6 prevails circumferential clearance. That is, in embodiments of FIG. 24, the asymmetrical arrangement of two first force application points 58, 60 with respect to the first return element 42 (left) is inverse to the asymmetric arrangement of two second force application points 58, 60 with respect to the second return element 42 (right) ,

It should also be pointed out here that the circumferential sequence of such different deflection mass pendulum units 72 does not necessarily have to be alternating, as indicated in FIG. 24. Also, in each case a plurality of deflection mass pendulum units 72 with basically the same construction, ie supporting functionality in the same circumferential direction, could be arranged consecutively. For reasons of symmetry and to avoid imbalance, however, the alternating arrangement shown in FIG. 24 is particularly advantageous.

In the mutual actuation (activation) of the play-bearing, but mutually prestressed restoring elements 42 according to FIG. 24, each restoring element 42 can be actuated alternately via the force application points 58, 60 located asymmetrically with respect to the restoring element 42 and thus used more effectively. The mutual bias of the pairwise return elements 42 also allows a play-free zero crossing with ideal characteristic curve, which takes the same course even when exceeding the bias voltage as in Fig. 25, curve c) (or Fig. 27, curve c), so that in total a characteristic curve according to FIG. 26 or 27, curve d) results. After passing through the game in the adjustment of the reset elements 42, which make a change of investment, the characteristic runs with the same slope as before, but offset parallel to the ideal characteristic of FIG 26 or 27, curve d). The load-free area can be determined here via the game to one of the two force application points 58, 60. In FIG. 26, curve d), the characteristic curve of this variant according to FIG. 24 is shown. Of particular advantage is the more effective use of the return elements 42 (mutually), which components can be saved, and the almost ideal characteristic curve. Compared to FIG. 15, curve a), the zero crossing here is free of play, as a result of which the sum stiffness of the torsional vibration damper is only minimally influenced at oscillation angles smaller pretensioning angle and at oscillation angles greater than the pretensioning angle. If the design is designed to be somewhat stiffer, the previously described stiffness losses can be compensated for by the clamping of the restoring elements 42 and the lower level of stiffness compared to the ideal stiffness in terms of sum stiffness.

In summary, all embodiments are characterized in that the position of at least one movable force application point on the centrifugal weight with respect to the cooperating restoring element in the rest position or at zero crossing is asymmetrical. That is, the radially extending return element 42, such as a leaf spring, can not be considered as the axis of symmetry of the force application points. In this case, the asymmetry can be achieved by an asymmetrical arrangement of the pin members 54, 56 in the flyweight 1 6, in such a way that in zero position a deflection of the return elements takes place. Alternatively, the return elements 42 can be placed asymmetrically with respect to the flyweight or flyweights (guideway of the flyweights) asymmetrically to the return elements 42. Likewise, by a combination of the various possibilities, a mutual distortion of the restoring elements 42 can be realized and tolerance-related play in the system can be eliminated. In embodiments, the first and the second restoring element 42 of a cooperating pair are arranged at least one, preferably exactly one, of their respectively associated force application points in each case without circumferential movement play. In other words, a restoring element 42 abuts in the rest position on a force application point cooperating therewith, so that the force application point or a bolt located behind it also acts as an abutment in the rest position without deflection of the deflection mass. A possibly existing second point of force application for the restoring element 42 is arranged therefrom on the other side with circumferential clearance (see FIG. 24).

27 summarizes a comparison of stiffness characteristics (force versus deflection for a rotational speed (DRZ)) of various embodiments. Curve a) relates to a backlash-free reciprocal actuation of the restoring elements according to FIG. 14. Curve b) relates to a two-sided mutual actuation of the restoring elements according to FIG. 13. Curve c) relates to a backlash-free, mutually braced and one-sided actuation of the restoring elements according to FIG. 23. The curve c-1 .5) relates to a backlash-free, mutually strained and one-sided control of the 1 .5 times the number of reset elements. The curve c-2.0) relates to a play-free, mutually strained and one-sided control of the 2.0 times the number of return elements with respect to the curve c). The curve d) relates to a play-related, mutually strained and reciprocal control of the restoring elements according to FIG. 24.

Hereinafter, with reference to FIGS. 28 to 32, various possible uses of the mutually preloaded torsional vibration damping arrangement 10 or 10 'described above will be explained.

In FIG. 28, a drive train 100 includes a drive unit 102 configured, for example, as an internal combustion engine. In the torque flow between the drive unit 102 and a transmission 104, for example an automatic transmission, a torsional vibration damping arrangement 10 in a rotating wet space 106 of a start-up element generally designated 108 is provided according to an embodiment. arranged game. This includes the stiffness provided by the deflection mass pendulum assemblies 72 with the deflection mass 34 and is coupled to the carrier 12 to a rotating component of the powertrain 100. In this case, in the rotating wet space 106, two serially effective torsional vibration dampers 1 10, 1 12 may be provided, each having a primary side and a secondary side and intervening damper springs over which the torque transmitted between the drive unit 102 and the gear 104 is passed. In the illustrated embodiment, a secondary side of the torsional vibration damper 1 10 is coupled to a primary side of the torsional vibration damper 1 12 for providing an intermediate mass or an intermediate element 1 14, to which the carrier 12 is connected. In the torque flow to the transmission 104 and a transmission output shaft 1 1 6 followed by a generally designated 1 18 Kardanwellenanordnung with respective flexible disks 120, 122 and an intermediate propeller shaft 124. On the output side, the propeller shaft 124 is coupled to an axle or differential 126. From this, the torque is transmitted to rims 128 and tires 130. In association with different transmission waves, such. As the transmission output shaft 1 1 6 a transmission shaft between the differential and the rims 128 and the rim 128 and the tire 130 are illustrated due to their inherent elasticity respective stiffnesses.

While FIG. 28 shows a drive train 100 installed longitudinally in the direction of travel, that is to say with a longitudinally oriented drive unit 102 and longitudinally oriented gear 104, FIG. 29 shows a drive train 100 with transversely mounted drive unit 102 or gear 104. Between these is, for example, a torsional vibration damper 132 in the form of a dual mass flywheel whose secondary side is coupled to a friction clutch, for example a dry friction clutch 134. A trained example also with torsional vibration damper clutch plate 136 transmits the torque further to the example designed as a manual transmission 104. On the secondary side of the torsional vibration damper or dual mass flywheel 132 of the carrier 12 of the torsional vibration damping arrangement 10 is coupled. On the output side follow on the transmission output shaft 1 1 6 and a differential 126 and the drive axle with its two rims 128 and tire 130. Again, with St respective stiffness of the drive shafts or wheels are illustrated.

FIG. 30 illustrates another example of a part of a drive train 100 with a hydrodynamic torque converter 1 50 following a drive unit 102 as a starting element 108. A pump wheel 138 is provided in the housing or rotating wet chamber 106 thereof or rotating therewith. This axially opposite a turbine wheel 140 is provided. Between the impeller 138 and the turbine 140 is located a generally designated 142 stator. Parallel to the hydrodynamic torque transmission path, which includes the fluid circulation between impeller, turbine wheel and stator, a torque transmission path via a lock-up clutch 144 can be set. On the lock-up clutch follow the two torsional vibration damper 1 10, 1 12, between which an intermediate mass 1 14 is formed. To this the turbine wheel 140 as well as the carrier 12 of the torsional vibration damping arrangement 10 is coupled. It should be noted that the recognizable for example in Fig. 30 torsional vibration damper may have a known structure with two shields and an intermediate central disc, wherein either the two cover plates or the central disc of the primary side and the respective other assembly is then assigned to the secondary side , In each such torsional vibration damper, one or more sets of springs may be parallel or serially, possibly also stepped, in order to be able to obtain a correspondingly graded damping characteristic.

Torsional vibrations or rotational irregularities introduced via the drive unit 102 into the input region of the hydrodynamic torque converter can initially be reduced or damped in the torsional vibration damper 1 located upstream in the torque flow when the lockup clutch 144 is engaged or torque-transmitting. The then still introduced into the intermediate mass 1 14 torsional vibrations can then be further mitigated or redeemed under the action of the coupled thereto torsional vibration damping arrangement 10 by appropriate design to a stimulating order. An even further filtering or vibration damping can then take place by the further downstream in the torque flow further torsional vibration damper 1 12. It goes without saying that various variations can be made here. Thus, for example, the turbine wheel 140 could be directly coupled to a transmission input shaft, ie the secondary side of the torsional vibration damper 1 1 2, whereby the inertia of a transmission input shaft is increased. This would mean that none of the two torsional vibration dampers 1 1 0, 1 1 2 would be effective in the range of effect of the hydrodynamics of the torque converter with disengaged lock-up clutch 144.

In a further variant, the turbine wheel 140 could provide the or part of the deflection mass 34. Thus, a Funktionsverschmelzung and thus a compact size can be ensured. Such a configuration would have the consequence that whenever the lock-up clutch 144 is disengaged and a torque is to be transmitted via the turbine wheel 140, the torsional vibration damping arrangement 10 is also used for torque transmission, in which case the design can be such that in In this state, the rotational angle limiting functionality of the bolts 40 and openings 36 is effective, so that the restoring elements 42 are not excessively loaded. When the lock-up clutch 144 is engaged, the turbine wheel is effective only as a deflection mass and also contributes to viscous damping due to the fluidic interaction.

Also, of course, the lock-up clutch 144 in the torque flow between the two torsional vibration dampers 1 1 0, 1 1 2 or even thereafter, it being ensured that the turbine 140 is coupled on the output side to the lock-up clutch 144. In a corresponding manner, of course, the carrier 1 2 of the torsional vibration damping arrangement 1 0 could be coupled to the primary side of the torsional vibration damper 1 1 0 or the secondary side of the torsional vibration damper 1 1 2.

FIG. 31 shows an embodiment variant of a drive train 100 in which the drive unit 1 02 transmits its torque via a dual-mass flywheel 1 32 integrated, for example, in a rotating wet space 1 06. At its secondary side, the torsional vibration damping arrangement 1 0 with its carrier 12 tethered. The torque flow is then followed by a starting element, for example a friction clutch 134.

In FIG. 32, a hydrodynamic torque converter 150 is shown in a structural design in partial longitudinal section. Its housing 152 provides the rotating wet space 106 and includes a drive-side housing shell 154 and a driven-side housing shell 156, which also forms a Pumpenradschale and carries on its inner side a plurality of circumferentially about the axis of rotation A successive Pumpenradschaufeln 158. The impeller 138 thus provided is the turbine wheel 140 with its turbine blades 1 60 axially opposite. Between the impeller 138 and the turbine 140 is the stator 142 with its Leitradschaufeln 162nd

The bridging clutch 144 comprises drive-side friction elements or disks 164 coupled to the drive-side housing shell 154 for rotation, and output-side friction elements or disks 1 68 coupled to a friction element support 1 66 for rotation. These can be transmitted by a clutch piston 170 for transmitting torque or for engaging the lockup clutch 144 pressed against each other. The torsional vibration damper 1 10 following in the torque flow onto the lockup clutch 144 and positioned radially outward comprises as a primary side a central disk element 172 coupled to the friction element carrier 66. Axially on either side thereof there are cover disk elements 174, 176, which with their radially outer area substantially the secondary side of the Provide torsional vibration damper 1 10. By damper springs 180 of the torsional vibration damper 1 10, a torque between the central disk element 172, so the primary side, and the cover disk elements 174, 176, so the secondary side, transmitted.

With their radially inner region, the cover disk elements 174, 176 form a secondary side of the radially inwardly positioned second torsional vibration damper 1 12. Axially between these cover disk elements fixedly connected to each other is another central disk element 182, which is essentially a secondary side the further torsional vibration damper 1 12 provides and is coupled by damper springs 184 with the cover plate elements 174, 176 for torque transmission.

The two cover disk elements 174, 176 essentially also provide the intermediate mass arrangement 14 to which, for example, by means of the two cover disk elements 174, 176, bolt 186, which is also firmly connected to one another, is coupled to the support 12 of a torsional vibration damping or absorber arrangement 10 constructed according to the invention. The flywheel 34 of the torsional vibration damping arrangement 10 comprises the two ground rings 30, 32 and the carrier disk 28 and lies axially substantially between the two radially staggered torsional vibration dampers 1 10, 1 12 and the turbine wheel 140. By the shaping of the mass ring 32 with radially inner Beveled contour of this, the turbine 140 can be positioned axially overlapping, so that an axially compact size is possible. The absorber assembly 10 may thus according to embodiments of the secondary side of at least one of the torque transmitting rotary or torsional vibration damper 1 10, 1 12 be coupled. The speed-adaptive vibration damper 10 is thus, generally speaking, an additional mass which can be coupled via a variable spring system to the drive system or at least one of the torsional vibration dampers 1 10, 1 12. In the connection of the torsional vibration damping arrangement 10 shown in FIG. 32 to the secondary side of a torsional vibration damper 110, 112, for example within a torque converter or a two-mass flywheel (not shown), the torsional vibration damping arrangement 10 can be constructed comparatively easily because the residual excitation at the location of the connection of the torsional vibration damping arrangement 10 behind the series-connected damper springs 180 and 184 can be comparatively small.

It can be seen that the support 12 is mounted rotatably radially inwardly via a bearing 188, for example plain bearings or rolling element bearings, on an output hub 190 of the torsional vibration damper arrangement 10 which is connected to the central disk 182. The turbine wheel 140 is also connected to this output hub 190 by toothing engagement for common rotation, so that the torque transmitted via the turbine wheel is bypassed, bypassing the two series-effective rotary shaft. Vibration damper 1 10, 1 12 is introduced into the output hub 190. Alternatively, as already explained above, the turbine wheel 140 could be coupled to the carrier 12 or, in general, the intermediate mass 1 14 or coupled to the deflection mass 34 in order to increase its mass moment of inertia.

The features disclosed in the foregoing description, the appended claims and the appended figures may be taken to be and effect both individually and in any combination for the realization of an embodiment in its various forms.

Although some aspects have been described in the context of a device, it will be understood that these aspects also constitute a description of the corresponding method, so that a block or a component of a device is also to be understood as a corresponding method step or as a feature of a method step. Similarly, aspects described in connection with or as a method step also represent a description of a corresponding block or detail or feature of a corresponding device.

The embodiments described above are merely illustrative of the principles of the present invention. It will be understood that modifications and variations of the arrangements and details described herein will be apparent to others of ordinary skill in the art. Therefore, it is intended that the invention be limited only by the scope of the appended claims and not by the specific details presented in the description and explanation of the embodiments herein.

Claims

claims

1 . Torsional vibration damping arrangement (1 0, 1 0 '), in particular for the drive train (1 00) of a vehicle comprising

a carrier assembly (1 2) rotatable about an axis of rotation (A);

a deflection mass (34) which is relatively movable in the circumferential direction relative to the carrier arrangement (1 2),

wherein the carrier assembly (1 2) and the deflection mass (34) are rotatably coupled relative to each other via a plurality of circumferentially arranged and radially extending resilient return members (42), wherein a return member (42) is movable radially in each case under centrifugal force and the force application point (58; 58 '; 60; 60') associated with the restoring element (42) is deformable,

characterized in that

a first return element (42) in the rest position of the torsional vibration damping arrangement (10; 1 0 ^' ) is biased in a first direction and that a second return element (42) is biased in the rest position in a second direction opposite to the first direction.

Second torsional vibration damping arrangement (1 0, 1 0 ') according to claim 1, characterized in that the opposite biasing forces resulting from the opposite biasing forces in the rest position are equal in magnitude.

3. torsional vibration damping arrangement (1 0, 1 0 ^' ) according to claim 1 or 2, characterized in that the first restoring element (42) in the rest position of the torsional vibration damping arrangement (1 0, 1 0') in the first direction by direct contact with a first In the rest position in the second direction, the second restoring element (42) is prestressed in the second direction by direct contact with a force application point (60, 60 ^' ) associated with the second restoring element (42) is.

4. torsional vibration damping arrangement (10; 10 ') according to any one of the preceding claims, characterized in that in each case at least one movable force application point (58; 58'; 60; 60 ') by a along the associated return element (42) movable in the radial direction flyweight ( 1 6, 1 6 ') is provided.

A torsional vibration damping arrangement (10; 10 ') according to any one of the preceding claims, characterized in that the first and second return elements (42) form a pair of return elements (42) and that along the first return element (42) in the radial direction movable first centrifugal weight (1 6, 1 6 ') and along the second restoring element (42) movable in the radial direction second centrifugal weight (1 6, 1 6') with the first and the second return element (42) on respective different sides of the return element (42) are in contact to bias the pair of return elements (42) in the rest position in opposite directions.

A torsional vibration damping arrangement (10; 10 ') according to one of claims 4 or 5, characterized in that the position of the at least one movable force application point (58; 58'; 60; 60 ') at the flyweight relative to the associated return element (42) in FIG the rest position is asymmetrical.

7. torsional vibration damping arrangement (10, 10 ') according to one of claims 4 to 6, characterized in that a centrifugal weight (1 6; 1 6') is acted upon in the rest position with a radially inwardly acting biasing force.

8. torsional vibration damping arrangement (10; 10 ') according to any one of the preceding claims, characterized in that each return element (42) is associated with exactly one in the radial direction movable force application point (58; 58 ^' ; 60; 60 '), wherein the respective force application points ( 58; 58 ';60; 60 ^' ) of the first and second return members (42) are disposed on respective different sides of the return members (42) to obtain the opposite biases.

9. torsional vibration damping arrangement (10; 10 ') according to one of claims 1 to 7, characterized in that each return element (42) are associated with two radially movable force application points (58; 60; 58 ^' ; 60 ^' ), wherein the two Force application points (58; 60; 58 ^' ; 60 ^' ) move radially on opposite sides of the respective restoring element (42) under centrifugal force and wherein the two points of application of force with respect to the respective restoring element (42) are arranged asymmetrically in the rest position to the opposite preloads receive.

Torsional vibration damping arrangement (10; 10 ') according to claim 6 or 9, characterized in that the asymmetrical arrangement of two first force application points (58; 60; 58'; 60 ') with respect to the first return element (42) inversely to the asymmetric arrangement of two second force application points (58; 60; 58 '; 60 ^' ) with respect to the second return element (42).

1 1. A torsional vibration damping arrangement (10; 10 ') according to any one of the preceding claims, characterized in that the first and the second restoring element (42) is arranged to at least one of their respective associated force application points (58; 58 ^' ; 60; 60 ^' ) each without circumferential play.

12. The torsional vibration damping arrangement (10; 10 ') according to one of the preceding claims, characterized in that the first and the second restoring element (42) form a pair of circumferentially juxtaposed restoring elements (42), and wherein the torsional vibration damping arrangement comprises a plurality of Having circumferentially arranged pairs.

13. torsional vibration damping arrangement (10; 10 ') according to any one of the preceding claims, characterized in that the restoring element (42) comprises a return spring, in particular a leaf spring or a bar spring, in particular with a linear force characteristic (K3).

14. torsional vibration damping arrangement (10; 10 ') according to any one of the preceding claims, characterized in that the return element (42) Be minus the deflection mass (34) and / or with respect to the carrier assembly (12) is fixed.

15. A drive train (100) for a vehicle, comprising at least one torsional vibration damping arrangement (10; 10 ') according to one of the preceding claims.

1 6. Drive train according to claim 15, characterized in that the drive train comprises a starting element, preferably a hydrodynamic torque converter (150) or fluid coupling or wet-running friction clutch or dry friction clutch, wherein the at least one torsional vibration damping arrangement (10; 10 ') in the region of the starting element is provided.

17. Drive train according to claim 15 or 1 6, characterized in that the drive train comprises at least one torsional vibration damper (1 10; 1 12) with a primary side and against the return action of a spring arrangement with respect to the primary side rotatable secondary side, wherein the carrier assembly (12) of at least one torsional vibration damping arrangement (10, 10 ') non-rotatably connected to the primary or secondary side of the at least one torsional vibration damper (1 10, 1 12) is connected.

18. Drive train according to claim 17, characterized in that the at least one torsional vibration damper (1 10, 1 12) has a first (180) and a second (184) spring arrangements, wherein an intermediate element (1 14) against the restoring action of the first spring arrangement ( 180) is rotatable relative to the primary side and the secondary side is rotatable against the return action of the second spring assembly (184) with respect to the intermediate member (1 14).

19. The drive train according to claim 18, wherein the carrier arrangement (12) of the at least one torsional vibration damping arrangement (10; 10 ') is non-rotatably connected to the intermediate element (1 14) of the torsional vibration damper (1 10; 1 12).